A power steering device equipped with a conventional flow control valve for a hydraulic pump is disclosed in Japanese Patent Application No. 115052/1995 filed by the assignee.
FIG. 7(A) shows a hydraulic pump in which a flow control valve FV disclosed in the Japanese application is integrally incorporated. A vane pump VP is used as the hydraulic pump.
The vane pump VP includes a housing H constructed of a pump body 10 and a cover 11. The housing H is formed with a shaft hole 12, in which a shaft 14 is rotatably supported through a bearing 13 arranged in the shaft hole 12. The shaft 14 functions as a drive shaft for a rotor 15 arranged in the pump body 10. The rotor 15 has a plurality of vanes 16 radially incorporated therein.
Also, the rotor 15 is mounted thereon with a cam ring 17 as shown in FIG. 7(B) which is viewed along line X--X of FIG. 7(A). The cam ring 17 has an inner surface formed into an elliptic shape. Driving of the shaft 14 permits rotation of the rotor 15, during which the vanes 16 are accessed to the inner surface of the cam ring 17. Thus, the vanes 16 are rotated while being intimately contacted with the cam ring 17 and permit chambers independent from each other to be defined therebetween.
When the chambers thus defined each are subject to a contraction stroke, hydraulic oil or fluid is discharged from a discharge port; whereas when each of the chambers is subject to an expansion stroke, hydraulic fluid is sucked thereinto.
The rotor 15 and cam ring 17 are commonly provided on a side surface thereof with a side plate 18. This results in a high-pressure chamber 19 being defined on a rear surface side of the side plate 18, so that a pump discharge pressure may be guided to the high-pressure chamber 19. A pressure of hydraulic fluid in the high-pressure chamber 19 forces the side plate 18 against the rotor 15 to keep loading balance.
The flow control valve VP, which will be described hereinafter, is arranged in a manner to be integral with the pump body 10 of the vane pump VP, so that a body of the flow control valve VP may also act as the body 10 of the vane pump VP.
The shaft 14 of the vane pump VP is connected to an engine (not shown), so that starting of the engine permits rotation of the rotor 15 connected to the shaft 14. Thus, an increase in rotational speed of the engine leads to an increase in fluid discharge rate at which hydraulic fluid is discharged from the vane pump VP.
The hydraulic fluid thus discharged from the vane pump VP, as shown in FIGS. 4 to 6, is guided through a pump port 4 to a pressure chamber 8a of the flow control valve PV and fed from an actuator port 20a to a power steering circuit PS through a feed passage.
At this time, such flowing of the hydraulic fluid discharged results in a pressure difference occurring between both sides of a variable constriction 3 arranged in the course or middle of the feed passage. A pressure on an upstream side acts on a left end surface of a spool 7 positioned on a side of the pressure chamber 8a and a pressure on a downstream side acts on a right end surface of the spool positioned on a side of a pilot chamber 8b through a pilot passage 29.
However, the spool 7 is not permitted to be moved in a right-hand direction unless thrust obtained by multiplying the pressure difference between both sides of the variable constriction 3 by a pressure receiving area of the spool 7 exceeds initial load of a spring 9 or unless a predetermined fluid discharge rate of the pump is obtained, resulting in a pump port 4 and a drain port 5 being kept isolated from each other. Therefore, all hydraulic fluid discharged from the pump is fed to the power steering circuit PS (an interval a of a characteristic line K shown in FIG. 4(B)).
When a rotational speed of the engine is increased to increase a fluid discharge rate, resulting in the pressure difference between both sides of the variable constriction 3 being increased to a predetermined level or above, the spool 7 is moved in a right-hand direction against the spring 9. Then, the spool 7 is stopped at a position at which the thrust described above and elastic force of the spring 9 are balanced with each other, to thereby permit the pump port 4 and drain port 5 to communicate with each other at a degree of opening corresponding to the position. This causes hydraulic fluid discharged from the pump to be returned thereto through the drain pump 5 depending on the degree of opening, so that the hydraulic fluid may be fed toward the power steering circuit PS at a maximum feed rate Q1 kept constant.
The maximum feed rate Q1 may be set on the basis of maximum power assisting force required.
A further increase in engine speed or rotational speed of the engine causes a decrease in fluid feed rate Q at which hydraulic fluid is fed to the power steering circuit PS for reasons described hereinafter. More particularly, an increase in fluid discharge rate of the pump causes a further increase in pressure difference between both sides of the variable constriction 3, leading to further movement of the spool 7 in the right-hand direction. Such movement of the spool 7 causes a diameter increased section 23a of a constriction member 23 to forcedly enter a constriction hole or orifice 22b, so that a degree of opening of the variable constriction 3 may be reduced. Also, the constriction member 23 is different in constriction effect thereof between when the diameter increased section 23a partially enter the orifice 22b and when the former entirely enters the latter. More specifically, an increase in degree at which the diameter increased section 23a enters the orifice 22b increases a pressure difference between both sides of the variable constriction 3, so that movement of the spool 7 may be increased to increase a degree of opening of the variable constriction 3 which affects a degree of communication between the pump port 4 and the drain port 5.
Thus, as indicated by the interval b of the characteristic line K in FIG. 4(B), hydraulic fluid is fed toward the power steering circuit PS at the maximum feed rate Q1 kept substantially constant until a rotational speed of the engine or an engine speed N reaches a predetermined level; whereas when the engine speed N exceeds the predetermined level, a flow rate at which fluid is fed to the power steering circuit PS is decreased, to thereby reduce the power steering force.
The engine speed N is substantially proportional to a velocity of the vehicle, to thereby permit application of the power assisting force corresponding to the velocity.
A maximum pressure applied to the power steering circuit PS is determined by a relief valve. More specifically, an increase in load pressure in the power steering circuit PS leads to an abnormal increase in pressure in the first pilot chamber 8b, which then acts on a ball poppet 33. When the pressure exceeds a relief set pressure determined by a spring 32, it forcibly opens the ball poppet 33 to permit the first pilot chamber 8b and drain port 5 to communicate with each other.
Such communication between the pilot chamber 8b and the drain port 5 causes fluid to flow through a pressure sensing hole 24, resulting in a pressure loss occurring through the hole 24. This leads to an abrupt reduction in pressure in the pilot chamber 8b, to thereby cause the spool 7 to be moved in a right-hand direction as shown in FIG. 5, resulting in a degree of opening of each of the pump port 4 and drain port 5 being increased to reduce a rate at which the pump feeds hydraulic fluid or a fluid feed rate of the pump.
Then, when a pressure in the power steering circuit PS is reduced to a level below the relief set pressure, the ball poppet 33 is set on a valve seat 34, so that a maximum pressure in the power steering circuit PS may be kept constant.
Also, the flow control valve FV, as shown in FIGS. 4 to 6, is so constructed that a piston 35 is arranged opposite to the spool 7 and has a change-over spool 36 incorporated therein.
Further, in the flow control valve FV, the spool 7 and piston 35 are arranged opposite to each other in the first pilot chamber 8b connected through the pilot passage 29 to a downstream side of the variable constriction 3 with the spring being interposedly arranged therebetween. This permits the piston 35 to be abutted against the spring 9 in the first pilot chamber 8b.
The piston 35 described above is formed on a central portion thereof with a flange 37, which functions to partition a cylinder hole 50 into a second pilot chamber 38 and a drain chamber 39. The drain chamber 39 is formed therein with a stepped portion 39a acting as a stopper, against which the flange 37 is abutted to prevent further movement of the piston 35.
The second pilot chamber 38 is arranged opposite to the first pilot chamber 8b with the piston 35 being interposed therebetween. Also, the drain chamber 39 is permitted to communicate with a tank passage (not shown) and kept from communicating with the first pilot chamber 8b. The piston 35 has one pressure receiving surface 35a arranged so as to face the first pilot chamber 8b and the other pressure surface 35b facing the second pilot chamber 38 and including a pressure receiving surface of the flange 37. The second pressure receiving surface 35b is formed so as to have a pressure receiving area larger than that of the first pressure receiving surface 35a.
The piston 35 is formed therein with a spool hole 40 in a manner to extend in an axial direction thereof. The spool hole 40 is so arranged that one end or a left end thereof is open to the first pilot chamber 8b and the other end or a right end thereof is closed. The spool thus formed has the change-over spool 36 slidably inserted thereinto, so that a pressure in the first pilot chamber 8b acts on a left end surface of the change-over spool 36.
Further, the spool hole 40 of the piston 35 is formed therein with an annular groove 41, which is arranged so as to communicate with the second pilot chamber 38 via a passage hole 42 of piston 35.
The change-over spool 36, as shown in FIG. 5(B), is formed thereon with two lands 43 and 44 in such a manner that an annular recess 45 is arranged therebetween. The right-hand land 44 of the change-over spool 36 has elastic force of a spring 46 applied thereto.
The annular recess 45 is arranged so as to constantly communicate with the drain chamber 39 through a passage 47 irrespective of a position of the piston 35 moved. Also, the change-over spool 36 is formed with a communication hole 48, to thereby permit a chamber in which a spring 49 is received to communicate with the drain chamber 39 through the annular recess 45. The change-over spool 36 constructed as described above is so operated that the land 43 interrupts communication between the first pilot chamber 8b and the annular groove 41 and permits the second pilot chamber 38 to communicate with the drain chamber 39 through the annular recess 45 and passage 47, when the change-over spool 36 is at a normal position shown in FIG. 5(A).
When the pump 1 is actuated, hydraulic fluid discharged from the pump 1, as described above, is guided through the pump port 4 to the pressure chamber 8a, as well as through the variable constriction 3 to the power steering circuit PS.
During non-steering, the power steering circuit PS is kept neutral, so that the hydraulic fluid is returned to the tank, resulting in a load pressure in the power steering circuit PS or a pressure on the downstream side of the variable constriction 3 being reduced. This keeps the pressure from exceeding a pressure set by the spring 46, so that the piston 35 is maintained at the normal position shown in FIG. 5(A). This results in initial load of the spring 9 in the first pilot chamber 8b being kept at a relatively reduced level.
Thus, the spool 7 is moved in the right-hand direction, until thrust obtained by multiplying a pressure difference between the pressure chamber 8a and the first pilot chamber 8b by a pressure receiving area of the spool 7 overcomes elastic force of the spring 9 in the first pilot chamber 8b, resulting in the spool 7 being balanced with load of the spring 9. Such movement of the spool 7 permits the pump port 4 to communicate with the drain port 5, so that the amount of fluid fed to the power steering circuit PS is reduced correspondingly.
When a maximum feed rate Q2 during the non-steering is set to be less than the above-described maximum feed rate Q1, energy loss during the non-steering requiring no assisting force may be significantly reduced.
During steering, when a pressure in the first pilot chamber 8b exceeds the set pressure determined by the spring 46, the change-over spool 36 is moved in the right-hand direction against elastic force of the spring 46, to thereby permit the first pilot chamber 8b and annular groove 41 to communicate with each other. The annular groove 41, as described above, is kept communicating with second pilot chamber 38 through the passage hole 42, resulting in communication between the first pilot chamber 8b and the second pilot chamber 38. This permits a pressure on the downstream side of the variable constriction 3 to be applied to each of the first pilot chamber 8b and second pilot chamber 38. Such application of the pressure on the downstream side of the variable constriction 3 to both pilot chambers 8b and 38 causes the piston 35 to be moved in the left-hand direction due to a difference in pressure receiving area between the pressure receiving surfaces 35a and 35b of the piston 35 as shown in FIG. 6(A). The maximum amount of movement of the piston 35 is regulated by abutment between the piston 35 and the stopper or stepped portion 39a.
Such movement of the piston 35 forcibly compresses the spring 9, resulting in load of the spring 9 being relatively increased. Such an increase in load of the spring 9 relatively reduces the amount of movement of the spool 7 which causes thrust based on a difference between a pressure in the pressure chamber 8a and that in the first pilot chamber 8b to be balanced with load of the spring 9, so that the amount of fluid fed from the pressure chamber 8a to the drain port 5 may be decreased. This results in a fluid feed rate at which fluid is fed to the power steering circuit PS being increased to the maximum feed rate Q1, so that flow characteristics indicated at the characteristic line K in FIG. 4(B) are obtained during the steering.
As will be noted from the above, the conventional flow control valve is so constructed that during the steering, the maximum feed rate Q1 is ensured to provide the power steering circuit PS with sufficient power and during the non-steering requiring no assisting force, the maximum feed rate Q2 of the pump 1 is reduced as compared with the maximum feed rate Q1 ensured in the steering, to thereby minimize energy loss.
In the prior art described above, when the operation is changed from the non-steering state to the steering state, a pressure in the first pilot chamber 8b is increased to cause the change-over spool 36 to be moved in the right-hand direction against the spring 46, resulting in the first pilot chamber 8b and second pilot chamber 38 communicating with each other.
However, at this time, the left end surface of the change-over spool 36 is caused to be open directly to the annular groove 41, so that the passage thereof is rapidly increased in area. This causes the change-over piston 35 to be rapidly moved in the left-hand direction, resulting in a feed pressure under which fluid is fed to the power steering circuit PS or power assisting force being abruptly varied, so that a driver of the vehicle has a feeling of disorder.
When the operation is changed from the steering state to the non-steering state, the passage between the second pilot chamber 38 and the first pilot chamber 8b is rapidly closed, followed by opening of the second pilot chamber 38 to the drain chamber 39, resulting in such problems as described above likewise occurring.
Also, in the prior art described above, the piston 35 is incorporated directly in a cylinder hole 50 formed in the body 10, followed by closing of the cylinder hole 50 with a plug 51. This requires incorporation of the piston 35 in the flow control valve during assembling thereof, resulting in the assembling being highly troublesome and costly. Further, for example, when it is desired to eliminate the piston 35 from the flow control valve after incorporation of the piston 35 into the flow control valve, it is required to remove the plug 51 and then eliminate the piston 35 therefrom. Thus, such elimination of the piston 35 is highly troublesome.
Moreover, in the prior art described above, the sleeve 49 pressedly fitted in the body 10 and the spool 7 guided into the body 10 are inserted into the body 10 from a side of the pressure chamber 8a. In order to ensure a space required for the insertion, it is required to arrange a constriction plate 22 formed with the constriction hole or orifice 22b separately from the constriction member 23. Thus, it is required to threadedly hold the constriction plate 22 indirectly in the body 10 by means of the plug 20 after incorporation of the sleeve 49 and spool 7 in the body 10.
Unfortunately, arrangement of the constriction plate 22 separately from the constriction member 23 causes misregistration between the constriction member 23 of the spool 7 and the orifice 22b, to thereby fail to provide the constriction member 23 and orifice with satisfactory coaxiality. This adversely affects a constriction effect or function of the variable constriction 3 constituted by a combination of the orifice 22b and constriction member 23.
The present invention has been made in view of the foregoing disadvantage of the prior art. It is an object of the present invention to provide a flow control valve for a hydraulic pump which is capable of keeping a driver from having a feeling of any disorder when it is used for a power steering apparatus, constructing a piston into a cartridge-type structure to facilitate incorporation thereof in the flow control valve, and permitting a variable constriction to exhibit a stable constriction function.